Hydraulic drive system for electrically-operated hydraulic work machine

ABSTRACT

An electrically-operated hydraulic work machine drives an actuator with a hydraulic pump driven by an electric motor and exercises load sensing control by controlling the rotation speed of the electric motor. The useful life of an electrical storage device, which is an electrical power source for the electric motor, is increased by suppressing the horsepower consumption of the hydraulic pump. This prolongs the operating time of the electrically-operated hydraulic work machine, and reduces the size of the electric motor. A controller exercises load sensing control over a variable displacement main pump by controlling the rotation speed of the electric motor, and provides the main pump with a torque control device that reduces the delivery rate of the main pump when the delivery pressure of the main pump increases, or provides the controller with a control algorithm that performs the same function as the torque control device.

TECHNICAL FIELD

The present invention relates to a hydraulic drive system for ahydraulic excavator or other electrically-operated hydraulic workmachine that performs various types of work by driving an actuator witha hydraulic pump driven by an electric motor. More specifically, thepresent invention relates to a load sensing hydraulic drive system forcontrolling the delivery rate of a hydraulic pump in such a manner thatthe delivery pressure of the hydraulic pump is higher than the highestload pressure by a predetermined pressure.

BACKGROUND ART

An electrically-operated hydraulic work machine, such as a hydraulicexcavator, that performs various types of work by driving an actuatorwith a hydraulic pump driven by an electric motor is described in PatentDocument 1. The electrically-operated hydraulic work machine describedin Patent Document 1 includes a fixed displacement hydraulic pump drivenby an electric motor, and exercises load sensing control by controllingthe rotation speed of the electric motor in such a manner that apressure difference is maintained constant between the delivery pressureof the hydraulic pump and the highest load pressure of a plurality ofhydraulic actuators.

PRIOR ART LITERATURE Patent Document

-   Patent Document 1: JP,A 2008-256037

SUMMARY OF THE INVENTION Problems to be Solved by the Invention

The hydraulic drive system described in Patent Document 1 can exerciseload sensing control by controlling the rotation speed of an electricmotor without using a variable displacement pump that provides complexflow control. Therefore, a load sensing system can be easily mounted,for instance, in a small-size hydraulic excavator.

However, the hydraulic drive system described in Patent Document 1 usesthe fixed displacement hydraulic pump. Therefore, when the deliverypressure of the hydraulic pump is maximized, the displacement of thehydraulic pump is at its maximum and remains unchanged. Hence, when therotation speed of the electric motor is controlled to its maximum leveldue to load sensing, the delivery rate of the hydraulic pump ismaximized so that the horsepower consumption of the hydraulic pumpincreases to a value indicated by the product of the maximum deliverypressure and the maximum delivery rate. As a result, the outputhorsepower of the electric motor increases to increase the electricalpower consumption. In this instance, the electrical power consumptionfor cooling the electric motor also increases, thereby increasing theamount of discharge from a battery (electrical storage device), which isan electrical power source for the electric motor. This causes a problemin which the battery rapidly becomes exhausted to shorten the operatingtime of the work machine.

Further, the output of the electric motor needs to be determined inconsideration of the maximum horsepower consumption of the hydraulicpump. This causes another problem in which an electric motor having ahigh output is required.

An object of the present invention is to provide a hydraulic drivesystem that is capable of not only increasing the operating time of anelectrically-operated hydraulic work machine by suppressing thehorsepower consumption of a hydraulic pump to increase the useful lifeof an electrical storage device, which is an electrical power source foran electric motor, but also reducing the size of the electric motor whenused for the electrically-operated hydraulic work machine that drives anactuator with the hydraulic pump driven by the electric motor andexercises load sensing control by controlling the rotation speed of theelectric motor.

Means for Solving the Problems

(1) In accomplishing the above object, according to an aspect of thepresent invention, there is provided a hydraulic drive system for anelectrically-operated hydraulic work machine. The work machine has anelectric motor, a hydraulic pump driven by the electric motor, aplurality of actuators driven by a hydraulic fluid discharged from thehydraulic pump, a plurality of flow control valves for controlling theflow rate of the hydraulic fluid supplied from the hydraulic pump to theactuators, and an electrical storage device for supplying electricalpower to the electric motor. The hydraulic drive system includes anelectric motor rotation speed control system and a torque controldevice. The electric motor rotation speed control system exercises loadsensing control to control the rotation speed of the hydraulic pump insuch a manner that the delivery pressure of the hydraulic pump is higherthan the highest load pressure of the actuators by a target differentialpressure. The torque control device exercises control to prevent anabsorption torque of the hydraulic pump from exceeding a predefinedmaximum torque by decreasing the delivery rate of the hydraulic pumpwhen the delivery pressure of the hydraulic pump increases.

As described above, the torque control device, which exercises controlto prevent the absorption torque of the hydraulic pump from exceedingthe predefined maximum torque by decreasing the delivery rate of thehydraulic pump when the delivery pressure of the hydraulic pumpincreases, is included in addition to the electric motor rotation speedcontrol system, which exercises load sensing control. Therefore, thehorsepower consumption of the hydraulic pump is suppressed to reduce theelectrical power consumption of the electric motor. This makes itpossible to increase the useful life of the electrical storage device,which is an electrical power source for the electric motor. As a result,the operating time of the electrically-operated hydraulic work machinecan be prolonged. Further, as the electrical power consumption of theelectric motor is reduced, it is possible to reduce the size of theelectric motor.

(2) According to another aspect of the present invention, there isprovided the hydraulic drive system as described in (1) above, whereinthe electric motor rotation speed control system includes a firstpressure sensor for detecting the delivery pressure of the hydraulicpump, a second pressure sensor for detecting the highest load pressure,an inverter for controlling the rotation speed of the electric motor,and a controller. The controller includes a load sensing controlcomputation section that computes a virtual displacement of thehydraulic pump, which increases or decreases depending on whether adifferential pressure deviation between the difference between thedelivery pressure of the hydraulic pump and the highest load pressureand a target LS differential pressure is positive or negative, inaccordance with the delivery pressure and the highest load pressure,which are detected by the first and second pressure sensors, and withthe target LS differential pressure, computes a target flow rate of thehydraulic pump by multiplying the virtual displacement by a referencerotation speed, and outputs a control command to the inverter for thepurpose of controlling the rotation speed of the electric motor in sucha manner that the delivery rate of the hydraulic pump agrees with thetarget flow rate.

As described above, a concept of the virtual displacement of thehydraulic pump is introduced into the load sensing control computationsection to determine the target flow rate of load sensing control andexercise load sensing control by controlling the rotation speed of theelectric motor. This makes it easy to improve the performance of loadsensing control based on electric motor rotation speed control (see (4)and (5) below).

(3) According to yet another aspect of the present invention, there isprovided the hydraulic drive system as described in (1) or (2) above,wherein the hydraulic pump is a variable displacement hydraulic pump;and wherein the torque control device is a regulator incorporated in thehydraulic pump.

Consequently, a smaller-size hydraulic pump can be used than when ahydraulic pump regulator is used to exercise load sensing control.

(4) According to still another aspect of the present invention, there isprovided the hydraulic drive system as described in (2) above, whereinthe hydraulic pump is a fixed displacement hydraulic pump; wherein thetorque control device is configured to exercise one function of thecontroller incorporated herein; and wherein the controller furtherincludes a torque limit control computation section that, in accordancewith the delivery pressure of the hydraulic pump, which is detected bythe first pressure sensor, computes a virtual displacement limit valuethat decreases with an increase in the delivery pressure of thehydraulic pump, and determines a new virtual displacement by selectingeither the virtual displacement computed by the load sensing controlcomputation section or the virtual displacement limit value, whicheveris smaller, and computes the target flow rate of the hydraulic pump bymultiplying the new virtual displacement by the reference rotationspeed.

Consequently, as the hydraulic pump is of a fixed displacement type, thesize of the hydraulic pump can be reduced to conserve space.

(5) According to an additional aspect of the present invention, there isprovided the hydraulic drive system as described in (2) or (4) above,further including an operating device that designates the referencerotation speed, wherein the controller sets the reference rotation speedin accordance with a designation signal from the operating device, andcomputes the target LS differential pressure and the target flow rate inaccordance with the reference rotation speed.

Consequently, when an operator manipulates the operating device toreduce the reference rotation speed, the target LS differential pressureand the target flow rate both decrease. As this reduces changes in therotation speed of the electric motor and decreases the rotation speed ofthe electric motor, an excellent micromanipulation capability isobtained.

Effect of the Invention

In an electrically-operated hydraulic work machine that not only drivesan actuator by driving a hydraulic pump with an electric motor, but alsoexercises load sensing control by controlling the rotation speed of theelectric motor, control is exercised to prevent the absorption torque ofthe hydraulic pump from exceeding a predefined maximum torque bydecreasing the delivery rate of the hydraulic pump when the deliverypressure of the hydraulic pump increases. This makes it possible tosuppress the horsepower consumption of the hydraulic pump, reduce theelectrical power consumption of the electric motor, and increase theuseful life of an electrical storage device that serves as an electricalpower source for the electric motor. As a result, the operating time ofthe electrically-operated hydraulic work machine can be prolonged.Further, as the electrical power consumption of the electric motor isreduced, it is possible to reduce the size of the electric motor.Moreover, the size of a cooling system for the electric motor can alsobe reduced because the size of the electric motor can be reduced.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a diagram illustrating the configuration of a hydraulic drivesystem according to a first embodiment of the present invention that isused for an electrically-operated hydraulic work machine.

FIG. 2 is a functional block diagram illustrating processes performed bya controller 50.

FIG. 3 is a diagram illustrating pump torque characteristics of a torquecontrol device (Pq characteristics (pump delivery pressure-pumpdisplacement characteristics)).

FIG. 4 is an external view of a hydraulic excavator in which thehydraulic drive system according to the first embodiment is mounted.

FIG. 5A is a diagram illustrating the horsepower characteristics of ahydraulic drive system that exercises load sensing control bycontrolling the rotation speed of an electric motor in a prior-artmanner.

FIG. 5B is a diagram illustrating the horsepower characteristics of thehydraulic drive system according to the first embodiment.

FIG. 6 is a diagram illustrating the configuration of the hydraulicdrive system according to a second embodiment of the present inventionthat is used for an electrically-operated hydraulic work machine.

FIG. 7 is a functional block diagram illustrating processes performed bythe controller.

FIG. 8 is a diagram illustrating the torque characteristics of a mainpump and characteristics (torque control characteristics) that simulatetorque control defined in a computation section.

MODE FOR CARRYING OUT THE INVENTION

Embodiments of the present invention will now be described withreference to the accompanying drawings.

First Embodiment Configuration

FIG. 1 is a diagram illustrating the configuration of a hydraulic drivesystem according to a first embodiment of the present invention that isused for an electrically-operated hydraulic work machine. The firstembodiment relates to a case where the present invention is applied tothe hydraulic drive system for a front swing type hydraulic excavator.

Referring to FIG. 1, the hydraulic drive system according to the presentembodiment includes an electric motor 1, a variable displacementhydraulic pump (hereinafter referred to as the main pump) 2, a fixeddisplacement pilot pump 30, a plurality of actuators 3 a, 3 b, 3 c, . .. , a control valve 4, a pilot hydraulic fluid source 38, and a gatelock valve 100. The main pump 2 and the fixed displacement pilot pump 30are driven by the electric motor 1. The actuators 3 a, 3 b, 3 c, . . .are driven by a hydraulic fluid discharged from the main pump 2. Thecontrol valve 4 is disposed between the main pump 2 and the actuators 3a, 3 b, 3 c, . . . . The pilot hydraulic fluid source 38 is connected tothe pilot pump 30 through a pilot hydraulic line 31 to generate a pilotprimary pressure in accordance with a fluid discharged from the pilotpump 30. The gate lock valve 100 is positioned downstream of the pilothydraulic fluid source 38 to serve as a safety valve that is operated bya gate lock lever 24.

The control valve 4 includes a second hydraulic fluid supply line 4 a(internal path), a plurality of closed-center flow control valves 6 a, 6b, 6 c, . . . , a plurality of pressure compensating valves 7 a, 7 b, 7c, . . . , a plurality of shuttle valves 9 a, 9 b, 9 c, . . . , a mainrelief valve 14, and an unloading valve 15. The second hydraulic fluidsupply line 4 a is connected to a first hydraulic fluid supply line 2 a(piping) to which the fluid discharged from the main pump 2 is supplied.The flow control valves 6 a, 6 b, 6 c, . . . are connected to hydrauliclines 8 a, 8 b, 8 c, . . . branched off from the second hydraulic fluidsupply line 4 a, and used to control the flow rate and direction of thehydraulic fluid to be supplied from the main pump 2 to the actuators 3a, 3 b, 3 c, . . . . The pressure compensating valves 7 a, 7 b, 7 c, . .. are connected to hydraulic lines 25 a, 25 b, 25 c, . . . , whichconnect a meter-in throttle section of the flow control valves 6 a, 6 b,6 c, . . . to a directional control section thereof, and used to controlthe downstream pressure of the meter-in throttle section of the flowcontrol valves 6 a, 6 b, 6 c, . . . until it is equal to a highest loadpressure (described later). The shuttle valves 9 a, 9 b, 9 c, . . .select the highest pressure (highest load pressure) from the loadpressures of the actuators 3 a, 3 b, 3 c, . . . , and output theselected highest pressure (highest load pressure) to a signal hydraulicline 27. The main relief valve 14 is connected to the second hydraulicfluid supply line 4 a to prevent the pressure in the second hydraulicfluid supply line 4 a (the delivery pressure of the main pump 2) fromexceeding a preselected pressure. The unloading valve 15 is connected tothe second hydraulic fluid supply line 4 a into which the fluiddischarged from the main pump 2 is introduced. When the deliverypressure of the main pump 2 is higher than a pressure obtained by addinga cracking pressure (the preselected pressure for a spring 15 a) to thehighest load pressure, the unloading valve 15 opens to return the fluiddischarged from the main pump 2 to a tank T, thereby limiting anincrease in the delivery pressure of the main pump 2.

The flow control valves 6 a, 6 b, 6 c, . . . have load ports 26 a, 26 b,26 c, . . . , respectively. When the flow control valves 6 a, 6 b, 6 c,. . . are in neutral position, the load ports 26 a, 26 b, 26 c, . . .communicate with the tank T and output a tank pressure as a loadpressure. When the flow control valves 6 a, 6 b, 6 c, . . . are shiftedfrom the neutral position to a left or right operating position (shown),the load ports 26 a, 26 b, 26 c, . . . communicate with the actuators 3a, 3 b, 3 c, . . . , respectively and output the load pressures of theactuators 3 a, 3 b, 3 c, . . . .

The shuttle valves 9 a, 9 b, 9 c, . . . are connected to the load ports26 a, 26 b, 26 c, . . . in a tournament manner, and form a highest loadpressure detection circuit together with the load ports 26 a, 26 b, 26c, . . . and the signal hydraulic line 27. In other words, the shuttlevalve 9 a selects either the pressure of the load port 26 a of the flowcontrol valve 6 a or the pressure of the load port 26 b of the flowcontrol valve 6 b, whichever is higher, and outputs the selectedpressure. The shuttle valve 9 b selects either the output pressure ofthe shuttle valve 9 b or the pressure of the load port 26 c of the flowcontrol valve 6 c, whichever is higher, and outputs the selectedpressure. The shuttle valve 9 c selects either the output pressure ofthe shuttle valve 9 b or the output pressure of another similar shuttlevalve (not shown), whichever is higher, and outputs the selectedpressure. The shuttle valve 9 c is a shuttle valve at a final stage. Theoutput pressure of the shuttle valve 9 c is output to the signalhydraulic line 27 as the highest load pressure. The highest loadpressure output to the signal hydraulic line 27 is introduced into thepressure compensating valves 7 a, 7 b, 7 c, . . . and the unloadingvalve 15 through signal hydraulic lines 27 a, 27 b, 27 c, . . . .

The pressure compensating valves 7 a, 7 b, 7 c, . . . include pressurereceivers 21 a, 21 b, 21 c, . . . , which operate in a closing directionand receive the highest load pressure from the shuttle valve 9 c throughthe signal hydraulic lines 27, 27 a, 27 b, 27 c, . . . , and pressurereceivers 22 a, 22 b, 22 c, . . . , which operate in an openingdirection and receive the downstream pressure of the meter-in throttlesection of the flow control valves 6 a, 6 b, 6 c, . . . . The pressurecompensating valves 7 a, 7 b, 7 c, . . . exercise control so that thedownstream pressure of the meter-in throttle section of the flow controlvalves 6 a, 6 b, 6 c, . . . is equal to the highest load pressure. As aresult, control is exercised so that the differential pressure acrossthe meter-in throttle section of the flow control valves 6 a, 6 b, 6 c,. . . is equal to the pressure difference between the delivery pressureof the main pump 2 and the highest load pressure.

The unloading valve 15 includes a spring 15 a, a pressure receiver 15 b,and a pressure receiver 15 c. The spring 15 a operates in a closingdirection and sets the cracking pressure Pun0 of the unloading valve 15.The pressure receiver 15 b operates in an opening direction and receivesthe pressure in the second hydraulic fluid supply line 4 a (the deliverypressure of the main pump 2). The pressure receiver 15 c operates in aclosing direction and receives the highest load pressure through thesignal hydraulic line 27. When the pressure in the hydraulic fluidsupply line 4 a is higher than a pressure obtained by adding thepreselected pressure Pun0 for the spring 15 a (cracking pressure) to thehighest load pressure, the unloading valve 15 opens, returns thehydraulic fluid in the hydraulic fluid supply line 4 a to the tank T,and exercises control so that the pressure in the hydraulic fluid supplyline 4 a (the delivery pressure of the main pump 2) is equal to apressure obtained by adding the preselected pressure for the spring 15 aand a pressure derived from the override characteristics of theunloading valve 15 to the highest load pressure. The overridecharacteristics of the unloading valve are such that the inlet pressureof the unloading valve, namely, the pressure in the hydraulic fluidsupply line 4 a, increases with an increase in the flow rate of thehydraulic fluid returning to the tank through the unloading valve. Inthis document, the pressure obtained by adding the preselected pressurefor the spring 15 a and the pressure derived from the overridecharacteristics of the unloading valve 15 to the highest load pressureis referred to as the unload pressure.

The actuators 3 a, 3 b, 3 c are, for example, a boom cylinder, an armcylinder, and a swing motor of a hydraulic excavator, respectively. Theflow control valves 6 a, 6 b, 6 c are, for example, a boom flow controlvalve, an arm flow control valve, and a swing flow control valve,respectively. For convenience of drawing, the other actuators, such as abucket cylinder, a swing cylinder, and a travel motor, and flow controlvalves related to these actuators are not shown.

The pilot hydraulic fluid source 38 is connected to the pilot hydraulicline 31 and provided with a pilot relief valve 32 that maintains aconstant pressure in the pilot hydraulic line 31. Manipulating the gatelock lever 24 can switch the gate lock valve 100 between a position forconnecting a pilot hydraulic line 31 a to the pilot hydraulic line 31and a position for connecting the pilot hydraulic line 31 a to the tankT.

The pilot hydraulic line 31 a is connected to control lever devices 122,123, 124 (see FIG. 4), which generate a command pilot pressure (commandsignal) for manipulating the flow control valves 6 a, 6 b, 6 c, . . . tooperate the associated actuators 3 a, 3 b, 3 c, . . . . When the gatelock lever 24 is switched into the position for connecting the pilothydraulic line 31 a to the pilot hydraulic line 31, the control leverdevices 122, 123, 124 regard the hydraulic pressure of the pilothydraulic fluid source 38 as a primary pressure and generate the commandpilot pressure (command signal) in accordance with the operation amountof each control lever. When, on the other hand, the gate lock valve 100is switched into the position for connecting the pilot hydraulic line 31a to the tank T, the control lever devices 122, 123, 124 are unable togenerate the command pilot pressure even if their control levers aremanipulated.

In addition to the elements described above, the hydraulic drive systemaccording to the present embodiment also includes a battery 70(electrical storage device), a chopper 61, an inverter 60, a referencerotation speed designation dial 51 (operating device), a pressure sensor40, a pressure sensor 41, and a controller 50. The battery 70 serves asan electrical power source for the electric motor 1. The chopper 61boosts the DC power of the battery 70. The inverter 60 converts the DCpower boosted by the chopper 61 to AC power and supplies the AC power tothe electric motor 1. The reference rotation speed designation dial 51is manipulated by an operator to designate the reference rotation speedof the electric motor 1. The pressure sensor 40 is connected to thehydraulic fluid supply line 4 a of the control valve 4 to detect thedelivery pressure of the main pump 2. The pressure sensor 41 isconnected to the signal hydraulic line 27 to detect the highest loadpressure. The controller 50 inputs a designation signal of the referencerotation speed designation dial 51 and detection signals of the pressuresensors 40, 41, and controls the inverter 60.

The chopper 61, the inverter 60, the reference rotation speeddesignation dial 51 (operating device), the pressure sensors 40, 41, andthe controller 50 form an electric motor rotation speed control systemthat exercises load sensing control by controlling the rotation speed ofthe electric motor 1 and that of the main pump 2 in such a manner thatthe delivery pressure of the main pump 2 is higher than the highest loadpressure of the actuators 3 a, 3 b, 3 c, . . . by a target differentialpressure.

FIG. 2 is a functional block diagram illustrating processes performed bythe controller 50.

The controller 50 includes computation sections 50 a-50 m to performvarious functions.

The computation sections 50 a, 50 b input the detection signals Vps,VPLmax of the pressure sensors 40, 41, respectively, and convert theinput signals to the delivery pressure Pps of the main pump 2 and thehighest load pressure PPLmax, respectively. Next, the computationsection 50 c determines the difference between the pressure Pps and thepressure PPLmax to calculate an actual load sensing differentialpressure PLS (=Pps−PPLmax). Next, the computation section 50 d convertsthe designation signal Vec of the reference rotation speed designationdial 51 to the reference rotation speed N0, and the computation section50 e converts the reference rotation speed N0 to a target LSdifferential pressure PGR.

The computation section 50 f calculates a differential pressuredeviation ΔP between the target LS differential pressure PGR and theactual load sensing differential pressure PLS. The computation section50 g calculates a change (increase/decrease) Δq in a virtualdisplacement q* of the main pump 2 from the differential pressuredeviation ΔP. The computation section 50 g is configured so that thevirtual displacement change Δq increases with an increase in thedifferential pressure deviation ΔP. Further, the virtual displacementchange Δq is calculated in such a manner that it is a positive valuewhen the differential pressure deviation ΔP is positive and is anegative value when the differential pressure deviation ΔP is negative.The computation section 50 h calculates a current virtual displacementq* by adding the virtual displacement change Δq to the virtualdisplacement q* prevailing one computation cycle earlier.

Here, the virtual displacement q* of the main pump 2 is a computeddisplacement value of the main pump 2 for controlling the rotation speedof the electric motor 1 in such a manner that the actual load sensingdifferential pressure PLS agrees with the target LS differentialpressure PGR.

The computation section 50 i performs a limiting process so that theobtained virtual displacement q* is within the range between a minimumdisplacement qmin and a maximum displacement qmax of the main pump 2(not smaller than the minimum displacement qmin and not greater than themaximum displacement qmax).

The computation section 50 j calculates a target flow rate Qd of themain pump 2 by multiplying the obtained virtual displacement q* by thereference rotation speed N0. The computation section 50 k calculates atarget rotation speed Nd of the main pump 2 by dividing the target flowrate Qd by the maximum displacement qmax of the main pump 2. Thecomputation section 50 m converts the target rotation speed Nd to acommand signal (voltage command) Vinv, which is a control command forthe inverter 60, and outputs the command signal Vinv to the inverter 60.

The computation sections 50 a-50 c, 50 f-50 h form a load sensingcontrol computation section. In accordance with the delivery pressurePps and the highest load pressure PPLmax, which are detected by thepressure sensors 41, 42, and with the target LS differential pressurePGR, the load sensing control computation section computes the virtualdisplacement q* of the main pump 2 that increases or decreases dependingon whether the differential pressure deviation ΔP between thedifferential pressure PLS, which is the difference between the deliverypressure of the main pump 2 and the highest load pressure, and thetarget LS differential pressure PGR is positive or negative.

The hydraulic drive system according to the present embodiment furtherincludes a torque control device 17 that exercises control to reduce thedisplacement of the main pump 2 in accordance with an increase in thedelivery pressure of the main pump 2 for the purpose of preventing anabsorption torque of the main pump 2 from exceeding a predefined maximumtorque. The torque control device 17 is a regulator that is integralwith the main pump 2 and provided with springs 17 b 1, 17 b 2 and atorque control tilt piston 17 a to which the fluid discharged from themain pump 2 is introduced through a hydraulic line 17 c.

FIG. 3 is a diagram illustrating pump torque characteristics of thetorque control device (Pq characteristics (pump delivery pressure-pumpdisplacement characteristics)). The horizontal axis represents thedelivery pressure of the main pump 2, and the vertical axis representsthe displacement of the main pump 2. TP0 is a characteristics curve ofthe maximum displacement of the main pump 2. TP1 and TP2 arecharacteristics curves of torque control defined by the springs 17 b 1,17 b 2. P0 is a predetermined pressure determined by the springs 17 b 1,17 b 2 (a pressure at which constant absorption torque control isinitiated).

When the delivery pressure of the main pump 2 is not higher than thepredetermined pressure P0, the torque control tilt piston 17 a of thetorque control device 17 does not operate, and the displacement of themain pump 2 is represented by the maximum displacement qmax on thecharacteristics curve TP0. When the delivery pressure of the main pump 2increases and exceeds the predetermined pressure P0, the torque controltilt piston 17 a of the torque control device 17 operates, and thedisplacement of the main pump 2 decreases along the characteristicscurves TP1, TP2 between the predetermined pressure P0 and a maximumdelivery pressure Pmax of the main pump 2 (a preselected pressure forthe main relief valve 14). As a result, control is exercised to maintainthe absorption torque of the main pump 2 (the product of the pumpdelivery pressure and displacement) at a substantially constant valuefor the purpose of preventing the absorption torque from exceeding themaximum torque (limit torque) TM on the characteristics curves TP1, TP2.In this document, the above-mentioned control scheme is referred to astorque limit control, and a control scheme based on characteristicsobtained when the displacement of the hydraulic pump is expressed interms of delivery rate is referred to as horsepower control. Themagnitude of the maximum torque TM can be freely set by selectingappropriate strengths of the springs 17 b 1, 17 b 2.

FIG. 4 is an external view of the hydraulic excavator in which thehydraulic drive system according to the present embodiment is mounted.

Referring to FIG. 4, the hydraulic excavator, which is well known as awork machine, includes an upper swing structure 300, a lower travelstructure 301, and a swing-type front work implement 302. The front workimplement 302 includes a boom 306, an arm 307, and a bucket 308. Theupper swing structure 300 can swing the lower travel structure 301 byrotating the swing motor 3 c shown in FIG. 1. A swing post 303 ismounted at the front of the upper swing structure 300. The front workimplement 302 is vertically movably mounted on the swing post 303. Theswing post 303 horizontally pivots with respect to the upper swingstructure 300 when a swing cylinder (not shown) extends or contracts.The boom 306, arm 307, and bucket 308 of the front work implement 302vertically pivot when the boom cylinder 3 a, the arm cylinder 3 b, andthe bucket cylinder 12 extend or contract. The lower travel structure301 is configured so that a blade 305, which vertically moves when ablade cylinder 304 extends or contracts, is mounted on a central frame.The lower travel structure 301 travels when travel motors 6, 8 rotate todrive left and right crawlers 310, 311. FIG. 1 shows only the boomcylinder 3 a, the arm cylinder 3 b, and the swing motor 3 c, and doesnot show the bucket cylinder 3 d, the left and right travel motors 3 f,3 g, the blade cylinder 3 h, and circuit elements thereof.

A cabin (cab) 313 is placed on the upper swing structure 300. The cabin313 incorporates a cab seat 121, the front/swing control lever devices122, 123 (only the device on the right side is shown in FIG. 4), thetravel control lever device 124, and the gate lock lever 24.

˜Operations˜

Operations of the present embodiment will now be described.

<When the Control Levers are in Neutral Position>

When all operating devices, including the control levers of the controllever devices 122, 123, 124, are in neutral position, all the flowcontrol valves 6 a, 6 b, 6 c, . . . are in neutral position. Therefore,the load ports 26 a, 26 b, 26 c, . . . of the actuators 3 a, 3 b, 3 c, .. . are connected to the tank so that the highest load pressure of theactuators 3 a, 3 b, 3 c, . . . , which is detected by the shuttle valves9 a, 9 b, 9 c, . . . , is equal to the tank pressure. The tank pressureis detected by the pressure sensor 41.

Meanwhile, the electric motor 1 drives the main pump 2 to supply ahydraulic fluid to the hydraulic fluid supply lines 2 a, 4 a. Thehydraulic fluid supply line 4 a is connected to the flow control valves6 a, 6 b, 6 c, . . . , to the main relief valve 14, and to the unloadingvalve 15. When all the control levers are in neutral position, the flowcontrol valves 6 a, 6 b, 6 c, . . . are closed so that the deliverypressure of the main pump 2 rises to a pressure obtained by adding thepressure derived from the override characteristics to the preselectedpressure for the spring 15 c of the unloading valve 15.

Here, the preselected pressure of the unloading valve 15 is maintainedconstant by the spring 15 a. The preselected pressure is higher than thetarget LS differential pressure PGR, which is calculated by thecomputation section 50 e when the reference rotation speed N0 ismaximized. If, for instance, the target LS differential pressure PGR is2 MPa, the preselected pressure for the spring 15 a is approximately 2.5MPa and the delivery pressure (unload pressure) of the main pump 2 isapproximately 2.5 MPa. The pressure sensor 40 connected to the hydraulicfluid supply line 4 a detects the delivery pressure of the main pump 2.The delivery pressure of the main pump 2 is designated by Pmin.

As mentioned earlier, the detection signal of the pressure sensor 40 isVps, and the detection signal of the pressure sensor 41 is VPLmax. Thecontroller 50 calculates the virtual displacement q* of the main pump 2in accordance with the detection signals Vps, VPLmax and with thedesignation signal Vec of the reference rotation speed designation dial51, and then calculates the target flow rate Qd by multiplying thevirtual displacement q* by the reference rotation speed N0. Further, thecontroller 50 calculates the target rotation speed Nd of the main pump 2by dividing the target flow rate Qd by the maximum displacement qmax ofthe main pump 2, converts the target rotation speed Nd to the commandsignal Vinv for the inverter 60, and outputs the command signal Vinv tothe inverter 60.

Here, as mentioned earlier, when all the control levers are in neutralposition, the highest load pressure is equal to the tank pressure andthe delivery pressure of the main pump 2 is higher than the target LSdifferential pressure PGR. Hence, as PLS=Pps−PPLmax=Pps>PGR, thedifferential pressure deviation ΔP (=PGR−PLS) computed in the controller50 is a negative value so that the virtual displacement q* of the mainpump 2 decreases. The minimum displacement qmin and the maximumdisplacement qmax are set in the computation section 50 i with respectto the virtual displacement q* so that the virtual displacement q*decreases to the minimum displacement qmin and is held at the minimumdisplacement qmin. Consequently, the target flow rate Qd decreases toits minimum value. Further, the target rotation speed Nd of the mainpump 2 and the command signal Vinv for the inverter 60 both decrease totheir minimum values. As a result, the rotation speed of the electricmotor 1 is held at its minimum value.

Meanwhile, the prevailing delivery pressure of the main pump 2 is Pminas mentioned earlier. As Pmin<P0, the torque control tilt piston 17 a ofthe torque control device 17 does not operate so that the displacementof the main pump 2 is at its maximum qmax. The resulting state isrepresented by point A in FIG. 3.

As described above, the displacement of the main pump 2 is maintained atthe maximum displacement qmax. However, as the rotation speed of theelectric motor 1 is held at its minimum value due to load sensingcontrol exercised by controlling the rotation speed of the electricmotor 1, the flow rate delivered by the main pump 2 is also held at itsminimum value.

Here, when the minimum rotation speed of the electric motor 1 is Nmin,the following equations are obtained:

Qd=qmin×N0=qmax×Nmin

Nmin=N0×(qmin/qmax)

In other words, when the resulting actual displacement of the main pump2 is q and the controlled rotation speed of the electric motor 1 is N(hereinafter simply referred to as the rotation speed N), the actualdisplacement q, the virtual displacement q*, and the rotation speed Nare expressed by the following equations:

q=qmax

q*=qmin

N=Nmin=N0×(qmin/qmax)

<Independent Boom Raising (Light Load)>

When the control lever of a boom control lever device, which is eitherthe control lever device 122 or the control lever device 123, is movedin a boom raising direction to perform a boom raising operation, a pilotpressure supplied from the pilot hydraulic line 31 is used as a sourcepressure so that a boom raising remote control valve (not shown) of theboom control lever device exerts the pilot pressure on an end facepressure receiver of the flow control valve 6 a. This moves the flowcontrol valve 6 a to the left indicated in the figure. The hydraulicfluid in a hydraulic fluid supply line 5 from the main pump 2 flowsthrough the flow control valve 6 a by way of the pressure compensatingvalve 7 a and is supplied to the bottom of the boom cylinder 3 a.

In the above instance, the load pressure of the boom cylinder 3 a isintroduced from the signal hydraulic line 27 to the pressure receiver 15c of the unloading valve 15 through the load port 26 a of the flowcontrol valve 6 a and through the shuttle valves 9 a, 9 b, 9 c. As theload pressure of the boom cylinder 3 a is introduced to the pressurereceiver 15 c of the unloading valve 15, the cracking pressure of theunloading valve 15 is set to a pressure obtained by adding the loadpressure to the preselected pressure for the spring 15 c so that thedelivery pressure of the main pump 2 rises to a pressure obtained byadding the load pressure and the preselected pressure for the spring 15c to the pressure derived from the override characteristics. Thepressure sensors 40, 41 detect the resulting delivery pressure of themain pump 2 and the highest load pressure.

As is the case where all the control levers are in neutral position, thecontroller 50 exercises so-called load sensing control based on theelectric motor 1 in accordance with processing functions depicted by thefunctional block diagram of FIG. 2 by controlling the rotation speed ofthe electric motor 1 by increasing or decreasing the command signal Vinvfor the inverter until the pressure in the second hydraulic fluid supplyline 4 a, that is, the delivery pressure of the main pump 2, is higherthan the highest load pressure by the target LS differential pressurePGR. The virtual displacement q* for the load sensing control increasesor decreases in accordance with the operation amount of a control lever(demanded flow rate) and varies from the minimum to the maximum due tothe limiting process performed by the computation section 50 i. As aresult, the rotation speed of the electric motor 1 (the rotation speedof the main pump 2) also varies from the minimum to the maximum inaccordance with the operation amount of a control lever (demanded flowrate).

Meanwhile, when the delivery pressure of the main pump 2 is Pb and Pb<P0due to light load, the torque control tilt piston 17 a of the torquecontrol device 17 does not operate so that the displacement of the mainpump 2 is at its maximum. An example of the resulting state isrepresented by point B in FIG. 3.

Here, the maximum rotation speed of the electric motor 1 is the rotationspeed prevailing when the virtual displacement q* is qmax. When themaximum rotation speed is Nmax, the following equations are obtained:

Qd=qmax×N0=qmax×Nmax

Nmax=N0

More specifically, the resulting actual displacement q of the main pump2, the virtual displacement q*, and the rotation speed N are expressedby the following equations:

q=qmax

qmin<q*≦qmax

Nmin<N≦Nmax

(Nmin<N≦N0)

<Independent Boom Raising (Heavy Load)>

When the load pressure of the boom cylinder 3 a rises to raise thedelivery pressure of the main pump 2 (the pressure in the hydraulicfluid supply line 5) to or above the predetermined pressure P0, which isdetermined by the springs 17 b 1, 17 b 2 of the torque control device17, the controller 50 uses the electric motor 1 to exercise load sensingcontrol in the same manner as described under <Independent boom raising(light load)>. In this instance, too, the virtual displacement q* forthe load sensing control increases or decreases in accordance with theoperation amount of a control lever (demanded flow rate) and varies fromthe minimum to the maximum, as is the case described under <Independentboom raising (light load)>. Further, the rotation speed of the electricmotor 1 (the rotation speed of the main pump 2) also varies from theminimum to the maximum in accordance with the operation amount of acontrol lever (demanded flow rate).

Meanwhile, as the delivery pressure of the main pump 2 is not lower thanthe predetermined pressure P0 in the above instance, the torque controltilt piston 17 a of the torque control device 17 operates so as todecrease the displacement of the main pump 2. Hence, so-called torquelimit control is exercised so that the displacement of the main pump 2decreases with an increase in the delivery pressure of the main pump 2.An example of the resulting state is represented by point C in FIG. 3.The delivery pressure of the main pump 2 is Pc (>P0) and thedisplacement thereof is qc.

Here, as mentioned earlier, the characteristics curves TP1, TP2 shown inFIG. 3 are set by the springs 17 b 1, 17 b 2. Therefore, the absorptiontorque of the main pump 2 (the product of pump delivery pressure anddisplacement), namely, the drive torque of the electric motor 1, iscontrolled not to exceed the maximum torque (limit torque) TM on thecharacteristics curves TP1, TP2.

More specifically, the actual displacement q of the main pump 2, thevirtual displacement q*, and the rotation speed N are expressed by thefollowing equations:

q=qc

qmin<q*≦qmax

Nmin<N≦Nmax

(Nmin<N≦N0)

<Independent Boom Raising (Relief State)>

When, for instance, the boom cylinder 3 a extends to reach its strokeend, the delivery pressure of the main pump 2 (the pressure in thesecond hydraulic fluid supply line 4 a) further rises to reach apreselected pressure for the relief valve 14. When the relief valve 14actuates, the pressure in the second hydraulic fluid supply line 4 a ismaintained at a level (so-called relief pressure−Pmax) preselected by aspring of the relief valve 14. Further, the load pressure of the boomcylinder 3 a is introduced into the signal hydraulic line 27 through theload port 26 a of the flow control valve 6 a. This load pressure isequal to the above-mentioned relief pressure. In other words, in theresulting state, the pressure in the second hydraulic fluid supply line4 a is equal to the pressure in the signal hydraulic line 27 and is alsoequal to the relief pressure set by the relief valve 14.

Moreover, the detection signal Vps concerning the pressure in the secondhydraulic fluid supply line 4 a, which is generated by the pressuresensor 40, and the detection signal VPLmax concerning the pressure inthe signal hydraulic line 27, which is generated by the pressure sensor41, are introduced into the controller 50. The pressures indicated bythese detection signals are equal to each other and also equal to therelief pressure set by the relief valve 14.

In the above instance, the controller 50 increases or decreases thevirtual displacement q* of the main pump 2 in such a manner that thepressure in the second hydraulic fluid supply line 4 a is higher thanthe pressure in the signal hydraulic line 27 by the target LSdifferential pressure PGR. In this case, as PLS=Pps−PLmax=0<PGR, ΔP(=PGR−PLS) is a positive value so that the virtual displacement q* ofthe main pump 2 increases. The minimum displacement qmin and the maximumdisplacement qmax are set in the computation section 50 i with respectto the virtual displacement q*. When, for instance, the boom cylinder 3a reaches its stroke end, the virtual displacement q* increases to themaximum displacement qmax and is held at the maximum displacement qmax.Therefore, the target flow rate Qd increases to its maximum value,thereby increasing the target rotation speed Nd of the main pump 2 andthe command signal Vinv for the inverter 60 to their maximum values,respectively. As a result, the rotation speed of the electric motor 1 isheld at the maximum value Nmax, which is equal to the reference rotationspeed N0.

Meanwhile, as the delivery pressure of the main pump 2 is not lower thanthe predetermined pressure P0 in the above instance as well, the torquecontrol tilt piston 17 a of the torque control device 17 operates toexercise torque limit control for the purpose of reducing thedisplacement of the main pump 2. The resulting state is represented bypoint D in FIG. 3. The displacement of the main pump 2 decreases to theminimum displacement qlimit−min due to torque limit control.

More specifically, the resulting actual displacement q of the main pump2, the virtual displacement q*, and the rotation speed N are expressedby the following equations:

q=qlimit−min

q*=qmax

N=Nmax=Nd

The above-described operations are performed when the boom ismanipulated. However, the same operations are also performed when thecontrol lever of a control lever device related to the arm 307 or otherwork element is manipulated.

˜Advantages˜

FIG. 5A is a diagram illustrating the horsepower characteristics of ahydraulic drive system that exercises load sensing control bycontrolling the rotation speed of an electric motor in a prior-artmanner. FIG. 5B is a diagram illustrating the horsepower characteristicsof the hydraulic drive system according to the present embodiment. It isassumed that the displacement (fixed) of a fixed displacement hydraulicpump in the prior-art hydraulic drive system is the same qmax as themaximum displacement of the main pump 2 according to the presentembodiment shown in FIG. 3.

The prior-art hydraulic drive system, which exercises load sensingcontrol by controlling the rotation speed of an electric motor in theprior-art manner, uses a fixed displacement hydraulic pump. Therefore,when the delivery pressure of the hydraulic pump is at its maximum Pmax,the displacement of the hydraulic pump remains at its maximum qmax.Hence, when load sensing control is exercised to maximize the rotationspeed of the electric motor, the delivery rate of the hydraulic pump isat its maximum Qmax so that the horsepower consumption of the hydraulicpump increases to a value that is the product of the maximum deliverypressure Pmax and the maximum delivery rate Qmax (shaded area of FIG.5A). As a result, the output horsepower of the electric motor increasesto HM*, which corresponds to the horsepower consumption of the hydraulicpump, thereby increasing the electrical power consumption of theelectric motor. In this instance, the electrical power consumption forcooling the electric motor also increases. This increases the amount ofdischarge from a battery (electrical storage device), which is anelectrical power source for the electric motor. This causes a problem inwhich the battery rapidly becomes exhausted to shorten the operatingtime of the work machine.

Further, the output of the electric motor needs to be determined inconsideration of the maximum horsepower consumption of the hydraulicpump. This causes another problem in which an electric motor having ahigh output is required.

The present embodiment, on the other hand, not only exercises loadsensing control by controlling the rotation speed of the electric motor,but also includes and uses the torque control device 17 in conjunctionwith the variable displacement main pump 2 and exercises control, asdescribed under <Independent boom raising (heavy load)> and <Independentboom raising (relief state)>, so that the absorption torque of the mainpump does not exceed the maximum torque TM when the delivery pressure ofthe main pump 2 rises. When torque limit control is exercised over themain pump 2 as described above, the absorption torque of the main pump 2is maintained at or below the maximum torque TM if the delivery pressureof the main pump 2 rises. Further, control is exercised so that thehorsepower consumption of the main pump 2 does not exceed maximumhorsepower HM, which is obtained by multiplying the maximum torque TM bythe prevailing rotation speed of the main pump 2. As a result, thehorsepower consumption of the main pump 2 is suppressed. Hence, theoutput horsepower of the electric motor 1 is reduced to HM to reduce itselectrical power consumption as compared to a case where load sensingcontrol is exercised by controlling the rotation speed of the electricmotor in the prior-art manner. This makes it possible to increase theuseful life of the battery 70 and prolong the operating time of theelectrically-operated hydraulic work machine. Moreover, as the outputhorsepower of the electric motor 1 is decreased, the size of theelectric motor 1 can be reduced.

In addition, the present embodiment introduces a concept of hydraulicpump virtual displacement q* into load sensing control computationsections 50 a-50 c, 50 f-50 h of the controller 50, determines thetarget flow rate Qd for load sensing control, and exercises load sensingcontrol by controlling the rotation speed of the electric motor 1. Thismakes it easy to improve the performance of load sensing control basedon rotation speed control of the electric motor 1.

For example, the controller 50 sets the reference rotation speed N0 inaccordance with the designation signal Vec of the reference rotationspeed designation dial 51, and calculates the target LS differentialpressure PGR and the target flow rate Qd in accordance with themagnitude of the reference rotation speed N0.

Consequently, when the operator manipulates the reference rotation speeddesignation dial 51 to reduce the reference rotation speed N0, thetarget LS differential pressure PGR and the target flow rate Qd bothdecrease. As this reduces changes in the rotation speed of the electricmotor 1 and decreases the rotation speed of the electric motor 1, anexcellent micromanipulation capability is obtained. Further, a controlalgorithm performing the same functionality as the torque control device17 can be easily incorporated into the controller 50 as described inconjunction with a second embodiment of the present invention.

Second Embodiment

FIG. 6 is a diagram illustrating the configuration of the hydraulicdrive system according to the second embodiment of the present inventionthat is used for an electrically-operated hydraulic work machine. Thesecond embodiment also relates to a case where the present invention isapplied to the hydraulic drive system for a front swing type hydraulicexcavator.

˜Configuration˜

Referring to FIG. 6, the hydraulic drive system according to the presentembodiment differs from the hydraulic drive system according to thefirst embodiment. More specifically, the hydraulic drive systemaccording to the present embodiment uses a main pump 2A, which is of afixed displacement type. The main pump 2A does not include the torquecontrol device 17 for horsepower control. Further, hydraulic drivesystem according to the present embodiment uses a controller 50A thathas a control function of simulating horsepower control of the main pump2A (the function of the torque control device).

FIG. 7 is a functional block diagram illustrating processes performed bythe controller 50A.

The controller 50A has a control block that includes computationsections 50 a-50 h. The computation sections 50 a-50 h compute thevirtual displacement q* of the main pump 2A. Computation sections 50 r,50 s are added to the above-described control block so as to reduce themaximum value of the virtual displacement q* in accordance with thedelivery pressure of the main pump 2A.

More specifically, the computation section 50 r has a table in whichtorque control simulation characteristics are defined. The deliverypressure Pps of the main pump 2A, which is converted by the computationsection 50 a, is input to the computation section 50 r. The computationsection 50 r references the table and calculates a virtual displacementlimit value (maximum virtual displacement) q*limit that corresponds tothe delivery pressure Pps of the main pump 2A.

FIG. 8 is a diagram illustrating the torque characteristics of the mainpump 2A and characteristics (torque control characteristics) thatsimulate torque control defined in the computation section 50 r.

As the main pump 2A is of a fixed displacement type, the displacement ofthe main pump 2A remains constant over the whole range of the deliverypressure of the main pump 2A and is equal to the maximum displacementqmax on the characteristics curve TP0.

The torque control characteristics defined in the computation section 50r are formed of characteristics corresponding to the maximumdisplacement characteristics curve TP0 of the main pump 2A, whichprevails when the delivery pressure of the main pump 2A is lower thanP0, and a constant torque curve TP4, which prevails when the deliverypressure of the main pump 2A is not lower than P0.

As described above, the torque control characteristics are defined inthe computation section 50 r. Therefore, when the delivery pressure Ppsof the main pump 2A is low so that Pps<P0, the computation section 50 rcomputes q*limit=qmax in accordance with the characteristics curve TP0.When the delivery pressure Pps of the main pump 2A rises so that Pps≧P0,the computation section 50 r computes q*limit=qlimit in accordance withthe constant torque curve TP4.

As described in conjunction with the first embodiment, the computationsection 50 h computes the virtual displacement q* for load sensingcontrol. The computation section 50 s selects either the virtualdisplacement q* for the load sensing control computed by the computationsection 50 h or the virtual displacement limit value q*limit determinedby the computation section 50 r, whichever is smaller, and outputs a newvirtual displacement q**. Here, a rule for selecting either one of thevirtual displacement q* for the load sensing control and the virtualdisplacement limit value q*limit (e.g., a rule for selecting the virtualdisplacement q* for the load sensing control) when they are equal shouldbe predefined. The selection of a small value by the computation section50 s corresponds to control for reducing the displacement by the torquecontrol device 17 according to the first embodiment in the event of anincrease in the delivery pressure of the main pump 2A.

The other processes (the processes performed by the computation sections50 a-50 h and the computation sections 50 i-50 m) are the same as thosedepicted in FIG. 2.

The computation sections 50 r, 50 s form a torque limit controlcomputation section that, in accordance with the delivery pressure Ppsof the main pump 2A, which is detected by the pressure sensor 40,computes the virtual displacement limit value q*limit that decreaseswith an increase in the delivery pressure Pps of the main pump 2A, anddetermine a new virtual displacement q** by selecting either the virtualdisplacement q* calculated by the load sensing control computationsection (computation sections 50 a-50 c, 50 f-50 h) or the virtualdisplacement limit value q*limit, whichever is smaller.

˜Operations˜

Operations of the present embodiment will now be described.

<When the Control Levers are in Neutral Position>

When all the operating devices, including the control levers of thecontrol lever devices 122, 123, 124, are in neutral position, thedelivery pressure of the main pump 2A is Pmin, which is equivalent tothe preselected pressure for the spring 15 c of the unloading valve 15,as described under <When the control levers are in neutral position> inconjunction with an exemplary operation according to the firstembodiment. The resulting state is represented by point A1 in FIG. 9. Inthis instance, as mentioned earlier, the differential pressure deviationΔP (=PGR−PLS) computed by the computation section 50 f of the controller50A is a negative value. Thus, the virtual displacement q* for loadsensing control decreases.

Meanwhile, the delivery pressure Pps of the main pump 2A, which isdetermined by the computation section 50 a of the controller 50A, isPmin, and Pps<P0 in the computation section 50 r. Therefore, qmax iscalculated as the virtual displacement limit value q*limit from thetorque control simulation characteristics.

Here, as q*≦q*limit, the computation section 50 s selects the virtualdisplacement q* for the load sensing control computed by the computationsection 50 h and outputs the selection as a new virtual displacementq**.

The subsequent processes to be performed are the same as those describedunder <When the control levers are in neutral position> in conjunctionwith the first embodiment.

Here, the virtual displacement q** decreases to the minimum displacementqmin due to the limiting process performed by the computation section 50i, thereby minimizing the target flow rate Qd, the target rotation speedNd of the main pump 2A, and the command signal Vinv for the inverter 60.This ensures that the rotation speed of the electric motor 1 and thedelivery rate of the main pump 2A are both held at their respectiveminimum values.

More specifically, the actual displacement q of the main pump 2A, thevirtual displacement q*, and the rotation speed N are expressed by thefollowing equations:

q=qmax(fixed)

q**=qmin

N=Nmin=N0×(qmin/qmax)

<Independent Boom Raising (Light Load)>

When the control lever of a boom control lever device, which is eitherthe control lever device 122 or the control lever device 123, is movedin a boom raising direction to perform a boom raising operation, thevirtual displacement q* for the load sensing control computed by thecontroller 50A increases or decreases in accordance with the operationamount of the control lever (demanded flow rate). If, in this instance,the delivery pressure of the main pump 2A is a pressure Pb representedby point B1 in FIG. 9, the delivery pressure Pps of the main pump 2A,which is determined by the computation section 50 a of the controller50A, is lower than P0. Thus, the computation section 50 r calculatesqmax as the virtual displacement limit value q*limit from the torquecontrol simulation characteristics (the characteristics curve TP0 inFIG. 9).

As q*≦q*limit in the above case, too, the computation section 50 sselects the virtual displacement q* for the load sensing controlcomputed by the computation section 50 h and outputs the selection as anew virtual displacement q**.

The subsequent processes to be performed are the same as those describedunder <Independent boom raising (light load)> in conjunction with thefirst embodiment.

Here, the virtual displacement q** increases or decreases in accordancewith the operation amount of a control lever (demanded flow rate) andvaries from the minimum to the maximum due to the limiting processperformed by the computation section 50 i. As a result, the rotationspeed of the electric motor 1 (the rotation speed of the main pump 2A)also varies from the minimum to the maximum in accordance with theoperation amount of the control lever (demanded flow rate).

More specifically, the resulting actual displacement q of the main pump2A, the virtual displacement q*, and the rotation speed N are expressedby the following equations:

q=qmax(fixed)

qmin<q**≦qmax

Nmin<N≦Nmax

(Nmin<N≦N0)

<Independent Boom Raising (Heavy Load)>

In a heavy-load state in which the load pressure of the boom cylinder 3a rises, the virtual displacement q* for the load sensing controlcomputed by the controller 50A also increases or decreases in accordancewith the operation amount of a control lever (demanded flow rate). If,in this instance, the delivery pressure of the main pump 2A is thepressure Pb represented by point C1 in FIG. 9, the delivery pressure Ppsof the main pump 2A, which is determined by the computation section 50 aof the controller 50A, is higher than P0. Thus, the computation section50 r calculates qlimit (<qmax) as the virtual displacement limit valueq*limit from the torque control simulation characteristics (the constanttorque curve TP4 in FIG. 9). The relevant position on the constanttorque curve TP4 is represented by point C2 in FIG. 9. At point C2,q*limit=qc.

The computation section 50 s selects either the virtual displacement q*or the virtual displacement limit value q*limit, whichever is smaller,and outputs the selection as a new virtual displacement q**. Morespecifically, the computation section 50 s selects q* when q*≦q*limit orselects q*limit when q*>q*limit, and outputs the selection as the newvirtual displacement q**.

Subsequent processes to be performed are the same as those describedunder <Independent boom raising (heavy load)> in conjunction with thefirst embodiment.

Here, the virtual displacement q** is limited to q*limit. Thus, thetarget flow rate Qd, the target rotation speed Nd of the main pump 2A,and the command signal Vinv for the inverter 60 are similarly limited tolimit the rotation speed of the electric motor 1.

As described above, the controller 50 has the same functionality as thetorque control device 17 according to the first embodiment and exercisescontrol to prevent the absorption torque of the main pump 2A fromexceeding the maximum torque (limit torque) TM.

If, in the above instance, the rotation speed corresponding to thevirtual displacement limit value q*limit is Nlimit, the actualdisplacement q of the main pump 2A, the virtual displacement q**, andthe rotation speed N are expressed by the following equations:

q=qmax(fixed)

qmin<q**≦qlimit

Nmin<N≦Nlimit

<Independent Boom Raising (Relief State)>

When, for instance, the boom cylinder 3 a extends to reach its strokeend, the delivery pressure of the main pump 2 is held at the reliefpressure Pmax with the highest load pressure being equal to the reliefpressure, as mentioned earlier. The resulting state is represented bypoint D1 in FIG. 9. In this instance, as mentioned earlier, thedifferential pressure deviation ΔP (=PGR−PLS) computed by thecomputation section 50 f of the controller 50A is a positive value.Thus, the virtual displacement q* for load sensing control increases.

Meanwhile, the delivery pressure Pps of the main pump 2A, which isdetermined by the computation section 50 a of the controller 50A, isPmax. Thus, the computation section 50 r calculates qlimit−min, which isat point D2 in FIG. 9, as the virtual displacement limit value q*limitfrom the torque control simulation characteristics (the constant torquecurve TP4 in FIG. 9). As q*>q*limit, the computation section 50 sselects the virtual displacement limit value q*limit computed by thecomputation section 50 r and outputs the selection as a new virtualdisplacement q**.

Subsequent processes to be performed are the same as those describedunder <Independent boom raising (relief state)>.

Here, the virtual displacement q** is limited to q*limit−min. Thus, thetarget flow rate Qd, the target rotation speed Nd of the main pump 2A,and the command signal Vinv for the inverter 60 are similarly limited tolimit the rotation speed of the electric motor 1.

Consequently, control is also exercised in the above instance so as toprevent the absorption torque of the main pump 2A from exceeding themaximum torque (limit torque) TM.

If, in the above instance, the rotation speed corresponding toq*limit−min is Nlimit−min, the actual displacement q of the main pump2A, the virtual displacement q**, and the rotation speed N are expressedby the following equations:

q=qmax(fixed)

q**=qlimit−min

N=Nlimit−min

The above-described operations are performed when the boom ismanipulated. However, the same operations are also performed when thecontrol lever of a control lever device related to the arm 307 or otherwork element is manipulated.

˜Advantages˜

As is the case with the first embodiment, the present embodimentexercises control to prevent the absorption torque of the main pump 2Afrom exceeding the maximum torque TM and prevent the horsepowerconsumption of the main pump 2A from exceeding the maximum horsepowerHM, which is obtained by multiplying the maximum torque TM by theprevailing rotation speed of the main pump 2A. As a result, thehorsepower consumption of the main pump 2A is suppressed. Hence, theoutput horsepower of the electric motor 1 is reduced to HM to reduce itselectrical power consumption as compared to a case where load sensingcontrol is exercised by controlling the rotation speed of the electricmotor in the prior-art manner. This makes it possible to increase theuseful life of the battery 70 and prolong the operating time of theelectrically-operated hydraulic work machine. Moreover, as the outputhorsepower of the electric motor 1 is decreased, the size of theelectric motor 1 can be reduced.

Further, as the main pump 2A is of a fixed displacement type, thepresent embodiment makes it possible to reduce the size of the main pump2A, thereby conserving space.

<Other>

The foregoing embodiments may be variously modified within the spiritand scope of the present invention. In the foregoing embodiments, thepressure compensating valves 7 a, 7 b, 7 c, . . . are of a postposedtype, positioned downstream of the meter-in throttle section of the flowcontrol valves 6 a, 6 b, 6 c, . . . , and used to control the downstreampressures of all the flow control valves 6 a, 6 b, 6 c, . . . at thesame maximum load pressure for the purpose of equalizing thedifferential pressures across the flow control valves 6 a, 6 b, 6 c, . .. . Alternatively, however, the pressure compensating valves 7 a, 7 b, 7c, . . . may be of a preposed type, positioned upstream of the meter-inthrottle section of the flow control valves 6 a, 6 b, 6 c, . . . , andused to control the differential pressure across the meter-in throttlesection at a preselected value.

Further, the foregoing embodiments have been described on the assumptionthat a hydraulic excavator is used as the work machine. However, evenwhen the present invention is applied to a construction machine (e.g., ahydraulic crane or a wheel excavator) other than a hydraulic excavator,the same advantages are obtained as far as it is a work machine thatdrives a plurality of actuators in accordance with a fluid dischargedfrom the main pump.

DESCRIPTION OF REFERENCE NUMERALS

-   1 Electric motor-   2, 2A Hydraulic pump (main pump)-   2 a First hydraulic fluid supply line-   3 a, 3 b, 3 c, . . . Actuator-   4 Control valve-   4 a Second hydraulic fluid supply line-   6 a, 6 b, 6 c, . . . Flow control valve-   7 a, 7 b, 7 c, . . . Pressure compensating valve-   8 a, 8 b, 8 c, . . . Hydraulic line-   9 a, 9 b, 9 c, . . . Shuttle valve-   14 Main relief valve-   15 Unloading valve-   15 a Spring-   15 b Pressure receiver operable in opening direction-   15 c Pressure receiver operable in closing direction-   17 Torque control device-   17 a Torque control tilt piston-   17 b 1, 17 b 2 Spring-   21 a, 21 b, 21 c, . . . Pressure receiver operable in closing    direction-   22 a, 22 b, 22 c, . . . Pressure receiver operable in opening    direction-   24 Gate lock lever-   25 a, 25 b, 25 c, . . . Hydraulic line-   26 a, 26 b, 26 c, . . . Load port-   27, 27 a, 27 b, 27 c, . . . Signal hydraulic line-   30 Pilot pump-   31, 31 a Pilot hydraulic line-   32 Pilot relief valve-   38 Pilot hydraulic fluid source-   40, 41 Pressure sensor-   50, 50A Controller-   50 a-50 m Computation section-   50 r, 50 s Computation section-   51 Reference rotation speed designation dial-   60 Inverter-   61 Chopper-   70 Battery-   100 Gate lock valve-   122, 123 Control lever device-   q* Virtual displacement-   q*limit Virtual displacement limit value-   TP1, TP2 Torque control characteristics curve-   TP4 Constant torque curve

1. A hydraulic drive system for an electrically-operated hydraulic workmachine, the work machine having an electric motor (1), a hydraulic pump(2) driven by the electric motor, a plurality of actuators (3 a-3 c)driven by a hydraulic fluid discharged from the hydraulic pump, aplurality of flow control valves (6 a-6 c) for controlling the flow rateof the hydraulic fluid supplied from the hydraulic pump to theactuators, and an electrical storage device (20) for supplyingelectrical power to the electric motor, the hydraulic drive systemcomprising: an electric motor rotation speed control system (40, 41, 50,51, 60, 61) that exercises load sensing control to control the rotationspeed of the hydraulic pump in such a manner that the delivery pressureof the hydraulic pump is higher than the highest load pressure of theactuators by a target differential pressure; and a torque control device(17; 50 r, 50 s) that exercises control to prevent an absorption torqueof the hydraulic pump from exceeding a predefined maximum torque bydecreasing the delivery rate of the hydraulic pump when the deliverypressure of the hydraulic pump increases.
 2. The hydraulic drive systemaccording to claim 1, wherein the electric motor rotation speed controlsystem includes a first pressure sensor (40) for detecting the deliverypressure of the hydraulic pump, a second pressure sensor (41) fordetecting the highest load pressure, an inverter (60) for controllingthe rotation speed of the electric motor, and a controller (50; 50A);wherein the controller includes a load sensing control computationsection (50 a-50 c, 50 f-50 h) that computes a virtual displacement (q*)of the hydraulic pump, which increases or decreases depending on whethera differential pressure deviation (ΔP) between the difference (PLS)between the delivery pressure (Pps) of the hydraulic pump (2; 2A) andthe highest load pressure (PPLmax) and a target LS differential pressure(PGR) is positive or negative, in accordance with the delivery pressureof the hydraulic pump and the highest load pressure, which are detectedby the first and second pressure sensors, and with the target LSdifferential pressure, computes a target flow rate (Qd) of the hydraulicpump by multiplying the virtual displacement by a reference rotationspeed (N0), and outputs a control command (Vinv) to the inverter for thepurpose of controlling the rotation speed of the electric motor (1) insuch a manner that the delivery rate of the hydraulic pump agrees withthe target flow rate.
 3. The hydraulic drive system according to claim1, wherein the hydraulic pump is a variable displacement hydraulic pump(2); and wherein the torque control device is a regulator (17)incorporated in the hydraulic pump (2).
 4. The hydraulic drive systemaccording to claim 2, wherein the hydraulic pump is a fixed displacementhydraulic pump (2A); wherein the torque control device is configured toexercise one function of the controller (50A) incorporated herein; andwherein the controller further includes a torque limit controlcomputation section (50 r, 50 s) that, in accordance with the deliverypressure (Pps) of the hydraulic pump, which is detected by the firstpressure sensor (40), computes a virtual displacement limit value(q*limit) that decreases with an increase in the delivery pressure ofthe hydraulic pump, and determines a new virtual displacement (q**) byselecting either the virtual displacement (q*) computed by the loadsensing control computation section (50 a-50 c, 50 f-50 h) or thevirtual displacement limit value, whichever is smaller, and computes thetarget flow rate (Qd) of the hydraulic pump by multiplying the newvirtual displacement by the reference rotation speed (N0).
 5. Thehydraulic drive system according to claim 2, further comprising: anoperating device (51) that designates the reference rotation speed (N0);wherein the controller (50, 50A) sets the reference rotation speed inaccordance with a designation signal from the operating device, andcomputes the target LS differential pressure (PGR) and the target flowrate (Qd) in accordance with the reference rotation speed.
 6. Thehydraulic drive system according to claim 2, wherein the hydraulic pumpis a variable displacement hydraulic pump (2); and wherein the torquecontrol device is a regulator (17) incorporated in the hydraulic pump(2).
 7. The hydraulic drive system according to claim 4, furthercomprising: an operating device (51) that designates the referencerotation speed (N0); wherein the controller (50, 50A) sets the referencerotation speed in accordance with a designation signal from theoperating device, and computes the target LS differential pressure (PGR)and the target flow rate (Qd) in accordance with the reference rotationspeed.